Stabilizing structure for railway car spring suspension



Feb. 17, 1959 s. a. GUINS 2,873,691

STABILIZING STRUCTURE FOR RAILWAY CAR SPRING SUSPENSION Filed July 24, 1956 4 Sheets-Sheet 1 FIG..1.

, Inventor SERGE! G. Gyms Feb. 17, 1959 s. G. GUINS 2,873,691

smsmzmc STRUCTURE FOR RAILWAY CAR SPRING SUSPENSION Filed July 24, 1956 4 SheetsSheet '2 a Q 6 Q 0 70 M.P,H, I) @S c 27 T5! ov' g *9 e0 MRH. 0 x0 R 5% 4- 4 35 1 o 1 w 3\ 34 50 MR H. v E 5 w d Z1 ,?J: 4o MRH. H

5 Fa! fr eauency of T, cal vLbr-atic ns 3.3 g +30 MP. H. 6 2 I q- 3. A fXtl-UQll eauenc -rvLL vibrations g a (I-Ughdenewiy 100. M.P. H, W 26 7 6 5'8 C5 Q E 1 Q 30 Natural freq uency-roll I0 MR H. 1 vibr tLom; low density load Load on car springs in thousands ofpounds INVEN TOR.

SERGEl G.GuINs S. G. GUINS Feb. 17, 1959 STABILIZING STRUCTURE FOR RAILWAY CAR SPRING SUSPENSION Filed July 24, 1956 4 Sheets-Sheet 4 Inventor SERGE! G. Gums United States Patent STABILIZING STRUCTURE FOR RAILWAY CAR SPRING SUSPENSION Sergei G. Guins, Olin'stead Falls, Ohio Application m 24, 1956, Serial No. 599,776 7 Claims. cries-197 This invention relates to improved means forstabilizing or damping the load supporting spring-suspension of a railway car so as to prevent the build-up therein of vi bration frequencies of objectionalamplitudes.

It is known that a railway car is bounced vertically when its wheels pass over the "rail joints of the trackway and that this bouncing action sets up forced vibrations 'in the load's'pring suspension assembly of the car. In order to minirnze the forced vibrations incident to the track construction the joints of the track rails are usually arranged so that the rail joints at one side of the track are midway of the joints at the other side of the track. When the said forced vibrations approximate the amplitude and fre uency of'the 'natur'al'vibr'atin'g frequency of the load spring suspension, harmonics develop therein which, in the absence of damping resistance, build up spring oscillation of objectional amplitudes. In addition'tothe'spring vibrationsincident to'the vertical bounce ofa car, thereare forced vibrations which are equally undesirable, 's'uch 'forexamp'le as the vibrations incident to the lateral swaying o'r'roll of the car body or the "fore and aft lurching of the car body. All ofthese shock waves and the forced vibrationfrequencies resulting therefromvary in relation to 'theweight oftheload,the center of "gravity as affected by the density of such-load, and the speed of operation. a

Prior to the present invention friction wedge means have been used to damp the springoscillations of railway freight cars. In some'instances the friction Wedges have exerted constant'friction to resist the spring vibration and in other instances-progressively increasing friction is provided'in relation to the'spring-defiection. However, these prior developments have not been "entirely satisfactory, since they have notprovided damping of sufiicient s'ensitivity to take into; account'the variations in damping requirements :for light and heavy loads and'for the lateral roll and other movements of the car body under such load variations. Itis'a principal objectof the present invention'toiprovid'e an improved spring stabiliiing structure which will provide adequate and "variable damping force for controlling the forcing frequencies of a railway car spring suspension in such manner as to take into'account' the difierent'frequenciesand consequently the different damping requirements foreont'r-oning'the spring oscillation resulting from the'combinedvertical movement, the lateral roll and other movements ofthe-ca'r bodyresultingfrom track and operating conditions.

It'has been found desirable to provide-greater proportion'ate dar'nping'force fortlie roll factor for-lighterloads than'is required underlikecnnditions forve'rtical motion only. This is true because th'e'naturalfrequenceis of'vi bration incident to such roll, particularly"for'lighter loads, come within theh'a'rmonic range. As the load increases, the natural frequency dueto such roll decreases and are below the'harmo'nic range for car weights-of twenty-thousand (20,000) pounds ormore Consequently these'vibrations become less'criticalwith increasing loads. How- 2,873,691 Patented Feb. 17, 1959 ever, the ratio of the natural frequency of vertical vibrations to the forcing frequency is close to the harmonic range at all operating speeds and under all load conditions and therefore requires damping to prevent objectional oscillations.

According to the present invention, the improved stabilizing structure includes the provision of spring actuated friction wedges interposed between relatively movable elements of a car truck, for example the truck bolsters and the truck side frames of a car; each bolster being resiliently supported on load springs carried on the truck side frames at opposite sides of the car.

The spring structures for actuating the friction wedges are constructed so that different portions of each spring has different degrees of stiffness whereby one or more turns or convolutions of the spring will be deflected or pressed together by predetermined loads and thereby vary the effective length of the spring and consequently vary the natural frequency rate of its vibration. The length of such spring, before any part thereof is bottomed or pressed together, provides the damping force for the lighter load, for example below twenty thousand (20,000) pounds, during the presence of spring vibrations induced by the combined vertical and roll movements of the car, and the convolutions or turnsof the springs which remain spaced apart above 20,000 pounds provide the damping force required for load conditions greater than twenty thousand (20,000) pounds.

The invention is illustrated in certain preferred embodiments in the accompanying drawings wherein:

'Fig. 1 is a fragmentary representation of a railway car truck provided with a spring stabilizing structure of the present invention;

'Fig. 2 is a transverse section taken substantially online 2-2 of Fig. 1, looking in the direction indicated by-the arrows;

Fig. 3 is a graphic representation of the natural frequencies of the spring suspension of a railway car resulting from the vertical and the'transverse rolling movement of the car body and illustratingalso the required damping -force for controlling the forced frequencies of said spring suspension under various service conditions;

Fig. 4 is one for'rn'of stabilizing spring structure constr 'ucted in accordance with the present invention;

Fig. 4a illustrates the sa'mespring when it is deflected by a load sulficient'to compress together certain convolutions which are formed to provide less stiffness than other convolutions of the spring;

Fig. 5 is another form of stabilizing spring structure made in accordance with the present invention;

Fig. 5 1 illustrates the relative deflections of the several convolutions ofthe spring under a predetermined 'load, for example'twenty thousand pounds;

Fig. '6 'is'a sectional view of a stabilizing spring structure composed of two concentrically arranged springs, the inner spring being longer and more flexible than the outer spring and is adapted to function independently of the outer spring for lighter loads and functions conjointly with the outer spring for heavier loads;

' 'Fig. 6a illustrates the spring structure'of Fig. 6 with onlythe inner-spring deflected by a predetermined load, mjxampk twenty thousand pounds; and

' Fig. 7 is a further modification ofa spring for actuating a friction Wedge of a structure for stabilizingthe spring suspension of a railway-car.

Referring first to Figs. land 2 of the drawings'wherein a'spring'stabilizing structure constructed accordingto the present'invention is illustrated in connection with cooperatingparts of'a'railway car truck: The said truck'and thecar-body supported thereon are of conventional constructiona'nd, for thesake of' 'brevity, areillustrated only fragmentarily. However, it will be understood that a ceiving an end portion 17 of the truck bolster. A group I of load springs 18 is carried on each side frame for resiliently supporting the truck bolster 12 in its operative position within the bolster windows 16 at opposite sides of the car.

The stabilizing structure for damping the forced vibrations imparted to the load spring 18 comprise wedge members 19-19 interposed between opposite sides of the truck bolster 12 and conventional wear plates 21-21 secured to forward and rear faces, respectively, of the bolster window 16. There are various approved arrangements for positioning the friction wedges. In some instances they are positioned to bear frictionally against the bolster 12 and in other instances they are positioned to bear frictionally against the vertical faces of the bolster window 16. In the present embodiment the wedges 19 are positioned within recesses 2222 formed in the opposite side faces of the bolster 12; each such recess being formed with a downwardly and inwardly inclined wall 23. The said wedges are each provided with an inclined face 24 which bears wedgingly against the inclined wall 23 of the recess in which the wedge is positioned. The said wedges are also formed with vertical faces 20 which bear frictionally against the wear plates 21. Each of the wedges is supported on a stabilizer spring 25 which seats on a lower portion of the truck side frame 13 and therefore functions to urge the associated wedge along the inclined wall 23 of the recess containing the wedge, whereby the cooperations of the inclined wall 23 and the spring force exerted by said stabilizing spring forces the wedge 19 laterally into frictional engagement with an adacent wear plate 21. Springs 25 are positioned beneath their associated wedges 19. Consequently, the frictional force exerted by the wedges is progressively increased in relation to the extent of downward movement of the bolster 12 under increasing weight of the car.

The designing of stabilizer springs, as previously constructed, took into account only the vertical frequencies incident to the vertical bounce of the car body and did not take into account the vibrations of the load springs incident to other movements of the car. The vertical vibrations of a railway car spring suspension represents a relatively simple movement and the stabilizing springs, as heretofore constructed, provided a force increasing uniformly with deflection. However, when the vertical movement of the car is combined with a lateral swaying or roll of the car body, the composite movement is highly complex and requires damping forces of diiferent degrees of resilience to satisfactorily meet the situation.

The problem of providing suitable damping force for controlling the combination of different vibration frequencies of a railway car spring suspension is graphically illustrated in Fig. 3 of the drawings. These illustrations were developed by the use of a basic force equation wherein F=48.2 ca/f and wherein: F represents the damping force required per each friction wedge of the spring stabilizing structure; 0 represents the percent of critical damping used; (1" represents the rate of acceleration in terms of gravity; and f represents the vibrating frequency.

The numerals along the margin at the left of Fig. 3 represent the vibrating frequencies of the load springs in cycles per second and also represent the required damping force, in thousands of pounds. The numerals along the bottom margin of said figure represents the load on the spring suspension, expressed in thousands of pounds, and the numerals along the margin at the right of said figure expresses the spe d of operation of t e ar in miles P 4 1 hour. The horizontal line 26 represents the forcing frequency imparted to the load springs of a railway car as an incident to the wheels of the car passing over the rail joints at the rate of 20 miles per hour; the rail joints at each side of the track being positioned midway of the joints at the other side of the track. The horizontal line 27 represents the forced frequencies which are induced by the normal track construction when the car is moving at the rate of 70 miles per hour. The curve 28 indicates the natural frequencies of vertical vibrations of the spring suspension of a railway car, and curves 29 and 30 represent the natural frequencies of the car springs incident to the roll of a car body when the car is empty or loaded, respectively, with the amount of load for each of the three curves 28, 29 and 30 being indicated by the numerals along the bottom margin of said Figure 3. By the use of said basic equation hereinabove referred to, the damping force required per each friction wedge 19 to eliminate vibration completely and known as critical damping force has been determined and the graph line 31 represents twenty percent (20%) of such critical damping force, since experience has demonstrated that twenty percent (20%) of the critical force is sufiicient for operating conditions. Accordingly, the load springs of an empty car weighing eight thousand (8,000) pounds have a forced vibrating frequency of approximately one (1) cycle per second (graph line 31) when the car is moving at approximately twelve and one-half (12%) miles per hour and a natural frequency (vertical) of 5.2 cycles per second at 70 miles per hour (graph line 28). This condition requires one thousand fifty (1,050) pounds of damping force to prevent objectionable amplification of the force induced frequencies. When the car is loaded to provide a weight of thirty six thousand (36,000) pounds (graph line BC) and operating at sixty (60) miles per hour, the forced vibrating frequency of the spring suspension is approximately four and one-half (4.5) cycles per second and a natural frequency of 2.9 cycles per second (graph line 28); thus the required damping force per each friction shoe is four thousand and five hundred (4,500) pounds. By the use of the said equation a graphic representation 32 is obtained for the damping force required, at twenty percent (20%) of the critical, for each friction wedge to control the roll vibration of a car carrying a high density load disregarding the range of forcing frequencies. A similar graph representation 33 indicates the damping requirements, under similar conditions, for a car carrying a low density load. However, both theory and experience indicate that high damping is required only when forcing frequencies approach natural frequencies. Graph 29 and 30 indicate that at loads above 20,000 lbs., forcing frequencies due to operating conditions are above natural frequencies, thus reducing need for damping force. Taking this in consideration graph 33 is adjusted and is illustrated in Figure 3 by graph line 34. This line intersects the graph line 31 at B and thereby gives as the resultant required damping force per shoe, an adjusted angular line A, B, C. It will be observed, therefore, that the damping force for the combination vibration incident to the vertical bounce and the roll of a car for loads up to twenty thousand (20,000) pounds is represented by the portion A, B of the graph line 34 and the damping force for controlling the vertical vibrations for loads above twenty thousand (20,000) pounds is represented by the portion B, C of the graph line 31.

The stabilizer spring 25, shown in Figs. 1, 2 and 4, is provided with an upper group of convolutions 35 designed to close if and when the car load is twenty thousand (20,000) pounds. Inasmuch as these convolutions are adapted to close at the lighter car loads makes the spring particularly suitable for damping the load spring vibrations induced by the combination vertical and roll movements of the car body for any weight of load, for example, by inspection of Fig. 3, it will be seen that 5. an empty car-weighing eight thousand (8,000 pounds will require damping force of approximately one thousand sevenhundred and fifty (1,750) pounds per shoe or three thousand and-five hundred (3,500) pounds for each group of load springs 18.

Assuming that the car is loaded so that the loadof the cargo plus the weight of the car is twenty thousand (20,000) pounds, the spacing'for the group of convolutions 35 of each stabilizer spring will close. In other words,-the portion 35 of eachspring functions to press its associated wedge 19; against the inclined wall 23 of the bolster and thereby, in cooperation with portion 36, to exert the damping forces represented by the portions A--B of the graph line 34. For heavier loads the portion 36 of the spring having the wider spacings between the convolutions thereof functions, in combination with said associated wedge, to provide the necessary damping forces represented by the portions BC of the graph line 31.

The portions 35 and 36 of the spring 25 function, in effect, as two springs arranged in axial alignment. For the purpose of simplifying the manufacture of a spring having the characteristics of the spring shown in Fig. 4, a modified spring structure as shown in Figs. 5, a is provided. The modified spring structure is designated generally by the reference numeral 25a. It is composed of an upper spring section 35a and a lower spring section 36a. The upper spring section may be made of steel of the same or different diameter than the lower spring section. For the purpose of illustration, the upper spring is made of stock slightly smaller in diameter than the stock of the spring shown in Fig. 4. The lower section 36a is constructed with spacings between the several convolutions to correspond substantially with the spacings between the convolutions of section 36 of Fig. 4. The two spring sections 35a, 36a are held in operative axial alignment by means of an adapter 37 comprising a central portion 38 and studs 39, 40 projecting outwardly from opposite sides of the central portion. The central portion 38 provides seating surfaces for the spring sections 35a, 36a and the studs 3940 fit into the central opening of the upper and lower spring sections, respectively.

Referring now to Figs. 6 and 6a wherein a stabilizing spring of the general character shown in Figs. 4 and 5 is composed of two springs 41, 42 arranged concentrically one within the other. The inner spring 41 is longer and more flexible than the outer spring 42, since it is designed to be deflected independently of the outer spring and to exert against a wedge 19 the force required to damp the spring suspension of a car when the load does not exceed a predetermined weight, for example twenty thousand (20,000) pounds. When the load exceeds the range for independent functioning of the inner spring 41 the inner spring is deflected sufliciently for the wedge 19 to engage the outer spring. Thereafter both springs 41 and 42 will function conjointly to exert the stabilizing force required for the heavier loads. For the lighter load conditions (below two thousand pounds) the forces exerted by the inner spring 41 follow the portion A--B of the graph line 31. For the heavier load conditions (above two thousand pounds) the forces exerted conjointly by the springs 41 and 42 follow the portion BC of graph line 31.

The further modification shown in Fig. 7 illustrates a stabilizing spring constructed of five groups of convolutions of different spacing. For example, groups of convolutions 49 and 50 are arranged at opposite ends of the spring and have spacings 51 and 52, respectively, adapted to close solid at a predetermined weight and are adapted to provide the spring pressure necessary against a friction wedge 19 to effect the stabilizing force desired for damping load spring vibrations of predetermined rate, for example spring vibrations induced by the fore and aft lurching movements of the car body. The

55, 56 between their respective convolutions of the groups and are designed to provide the necessary damping force for the vibrations of the load springs induced by the vertical androlling movements of the car body carrying the lighter loads, for example up to twenty thousand (20,000) pounds. An intermediate group 57 is provided with suitable spacings 58 between its several convolutions so that the groupwill provide the damping force required for theload springs when the latter are subjected to weights above twenty thousand (20,000) pounds.

It will be seen from the description of the invention that the improved stabilizing structure shown and described makes it possible to impart smooth riding qualities to a railway car provided with standard load springs in which the force exerted is increased uniformly with a deflection, since the improved stabilizing structure imparts variable damping forces to the load springs in opposition to the forced vibrations induced by a different combination of movements of the car body such, for example, as the vertical bouncing movements, the lateral or rolling movements, the longitudinal lurching and weaving movements, all of which may have different frequencies of vibration for different operating conditions.

I claim:

1. In combination with a railway car spring suspension including a group of load supporting springs, a first truck element on which said spring group is supported and a second truck element movable relative to said first element and supported on said spring group for transmitting the load weight of the car to said spring group, of a stabilizing structure for applying variable frictional forces to one of said elements to damp vibrations of variable frequencies induced in said load springs by various movements of the railway car under variou load weights and operating conditions comprising a friction wedge interposed between said first and second truck elements, and means for forcing said wedge into wedging engagement with one of said truck elements and into frictional engagement with the other comprising a stabilizer spring structure seated at one end on said first truck element with its other end engaging said wedge with resilient pressure; the said stabilizing spring structure having a group of convolutions having one rate of deflection for damping oscillations induced in the load springs below their natural frequency of vertical vibrations and another group having a different rate of deflection for damping oscillations induced in said load springs above their natural rate of vertical vibrations.

2. A combination structure according to claim 1 in which said groups of convolutions are separate springs assembled in axial alignment.

3. A combination structure according to claim 1 in which the groups of convolutions having different rates of deflection are constructed of stock of dilferent diameters.

4. A combination structure according to claim 1 in which the group of convolutions for stabilizing the oscillations induced in the load springs below their natural frequency of vertical vibration have a smaller pitch than the other group and function to apply the required forces for damping oscillations induced in the load springs by rolling movements of the railway car at lighter loads and the said another group of convolutions have a greater pitch for applying the forces required for damping oscillations induced by vertical bouncing of the car at heavier loads.

5. A combination structure as defined in claim 1 in which the first truck element is a side frame of the truck and in which the second mentioned truck element is a truck bolster.

6. A combination structure as defined in claim 1 characterized in that the stabilizing spring structure comprises two springs concentrically arranged one within the other,

the inner spring being longer and more flexible than the outer spring.

7. A combination structure as defined in claim 6 in which the inner spring is deflected independently of the outer spring for load conditions up to a predetermined weight and in which both springs function conjointly for load conditions above said predetermined weight.

References Cited in the file of this patent UNITED STATES PATENTS 190,061 Middleton Apr. 24, 1877 338,267 Hearle Mar. 23, 1886 8 Pfingst NOV. Pfingst Nov. Stone Nov. Graham, Nov. Samuels May. Clark Oct. Holland Oct. Light Dec. Barber Oct.

Wulfi Jan.

FOREIGN PATENTS France Apr. 

